An axial piston machine is a machine in which a plurality of axially extending cylinders, together comprising the cylinder cluster, is arranged in a generally rotationally symmetrical layout around a central axis coincident with the rotational axis of a crankshaft. Each cylinder contains a reciprocating piston and may reciprocate along an axis parallel or slightly inclined to that of the other cylinders. Axial piston machines may offer a number of potential advantages over other multi-cylinder piston machine configurations including: reductions in size and weight, simplified fluid porting, and the ability to achieve close to perfect balancing of the dynamic inertial forces.
There are a number of different mechanisms that can be used to drive the reciprocating motion of the pistons in their cylinders, two of the most common types being Swashplate drives and Z-Crank drives. While terminology can vary, a swashplate is in effect a cam surface attached to and rotating with the crankshaft that drives or is driven by the reciprocating linear motion of the pistons. Each piston has a bearing or bearings attached to it that slides or rolls over the surface of the swashplate cam surface. Each piston also has some form of linear bearing such as the side of the piston within its cylinder that reacts the lateral forces created by the action of the piston-driving bearings when on the inclined surface of the swashplate. The piston-swashplate bearings may have a sliding or rolling speed over the swashplate in the order of two times the peak piston speed. While this arrangement is adequate for axial piston machines having relatively low piston speeds such as compressors and hydraulic pumps or motors, modern internal combustion engines commonly have much higher piston speeds. Also inertial loads and bearing sliding or rolling speeds in a swashplate drive operating can lead to high frictional losses for the higher speed combustion engines making standard swashplate configurations less attractive for internal combustion engines.
Z-Crank drives employ an intermediate body known variously as a Wobbleplate, Wabbler, Reciprocator or Spider that rotates on reciprocator bearings. Such is mounted on and for rotation about a crank section of the crankshaft by reciprocator bearings. The inclined crank section has an inclined crank axis that is at an acute angle (hereinafter referred to as the “swash angle”), to and intersecting with the crankshaft's rotational axis. The intersection point is hereinafter referred to as “point X”.
The reciprocator is restrained against rotation relative to the cylinder cluster. Rotation of the inclined crank section by rotation of the crankshaft causes the reciprocator to nutate. As a result, points on the body of the reciprocator in a plane passing through point X and perpendicular to the crank axis, move in a predominantly axial oscillatory motion parallel to the crank shaft axis, with motion in the plane perpendicular to the crankshaft axis being of relatively small magnitude. Such points define the preferred location for engaging connection rods for the transmission of motion between a piston and the reciprocator.
The connection between the reciprocator and pistons can take many forms but generally connection rods, having sufficient rotational degrees of freedom at either end are utilised. The reciprocator bearings typically operate at much lower sliding speeds than would the swash plate piston bearings of an equivalent Swashplate drive. As a consequence frictional losses will generally be reduced and higher operating speeds may be made possible.
An important element of an axial piston machine design incorporating a Z-crank drive is the method used to restrain the reciprocator from rotation relative to the cylinder cluster (herein after referred to as the method of “rotation restraint”). Without such a restraint the reciprocator will generally not translate the rotation of the crankshaft into the necessary reciprocating motion of the pistons as desired. Depending on swash angle, the rotation restraint arrangement generally transmits a rotation similar in magnitude to the rotation delivered (or absorbed in the case of a pump or compressor) by the crankshaft.
A number of rotation restraint systems that have been employed.
U.S. Pat. No. 4,491,057 utilises a Universal joint, also known as a Cardan or Hooke's joint, to provide the rotation restraint for the reciprocator. The Universal joint is not a constant-velocity joint and as the crankshaft rotates the reciprocator is subjected to gimbal error that produces unbalanced angular accelerations and inertial rotations at twice the frequency of the crankshaft rotation. These accelerations and rotations increase greatly if the swash angle is increased and also become more pronounced at high speeds. The gimbal error has a period of 180 degrees of rotation of the cranks shaft relative to the reciprocator. So that for connection rods connected at other than 180 degree spacings about the cranks axis (i.e. of an axial piston machine with two pistons), the reciprocating motion of the pistons in their cylinders varies such that different pistons will not share the same displacement, velocity and acceleration cycles. The variation in fluid and thermodynamic processes between cylinders that this can lead to is generally undesirable. Therefore Cardan joints do not readily lend themselves to use in machines with odd numbers of pistons.
Another system that has been employed such as described in U.S. Pat. No. 6,003,480 or U.S. Pat. No. 4,852,418 uses a planar sliding guide, groove or cam surface attached to the cylinder cluster against which a complementary bearing attached to the reciprocator runs in order to provide reciprocator rotation restraint. Because the motion of the bearing on the reciprocator is held planar with respect to the cylinder cluster, the pistons in a Z-crank machine employing such a sliding rotation restraint system may be subjected to similar variations in the motions of the pistons as with universal joint rotation restraint systems. There may also be significant frictional losses associated with such a sliding rotation restraint system owing to the relatively high sliding velocities at the rotation restraint bearing contact point.
U.S. Pat. No. 5,094,195 utilises meshing bevel gears in which a reciprocator mounted bevel gear concentric to the reciprocator axis with the vertex of its conical teeth coincident with point X engages with a second identical bevel gear mounted off of the cylinder cluster concentric to the axis of the crankshaft with the vertex of its conical teeth also coincident with point X. The two bevel gears have the same number of teeth and the same cone angle equal to 180 degrees minus the swash angle. This bevel gear rotation restraint method has a number of possible disadvantages:                The bevel gear mounted on the reciprocator can add significantly to the mass of the reciprocator and contribute to higher inertial loadings on the reciprocator bearings.        The bevel gears engage at high speed and can be the source of significant frictional losses.        Bevel gears subjected to the pulsating rotations generated by internal combustion or gas compression can also be subjected to significant impulsive loads that may require heavier gears and can also generate significant noise as a result of backlash.        Bevel gears generally requite precise alignment to run quietly and efficiently without wear, and this can be difficult to achieve within the highly loaded dynamic environment of the reciprocator.        Bevel gears are generally limited to operating at one fixed swash angle.        Space and geometrical restraints can make it difficult to build sufficiently robust bevel gears into the machine given the necessity placement of other components and the rotation transmitted. Placing suitable bevel gears in the required locations about point X could also compromise the design of the reciprocator, the conical face of the bevel gears generally needs to lie either radially inside or outside of the circular array of connecting rods, the inner radial location can lead to structural compromises in design that may increase reciprocator mass and reciprocator bearing frictional losses. The outer radial location for a bevel gear on the reciprocator may have generally fewer structural compromises, but may increase the overall diameter of the engine significantly. A large diameter bevel gear on the periphery of the reciprocator will generally introduce a lot of reciprocating mass with consequently higher reciprocator bearing loads. Higher pitch-line-engagement velocities for large diameter bevel gears peripheral to the reciprocator and conrods may lead to unacceptable noise, friction and wear.        
U.S. Pat. No. 5,450,823 employs a homo-kinetic or Constant Velocity (CV) type joint in the form of a double Universal joint, incorporating two Universal joints connected together through a short intermediate shaft such that the gimbal error from each joint is cancelled out by the other. There is a significant degree of complexity and a large number of bearings involved in the arrangement and the required placement of this rotation restraint mechanism may also make the reciprocator heavier due to the less ideal load paths in the reciprocator that is built around it.
Another solution that has been suggested for rotation restraint in U.S. Pat. No. 5,129,752 is to employ Ball-and-Crevice or Rzeppa type constant velocity (CV) joints as are commonly used to drive the front wheels of cars. In most CV applications the balls in such a joint orbit in a circular path. They hence have inertial forces that are principally centrifugal and are reacted by the enclosing housing with little friction, but in a Z-crank machine the balls are continuously accelerated back and forth in harmonic motion along an arcuate path. These alternating accelerations and decelerations at the relatively high speeds of an axial piston internal combustion engine in combination with impulsive and even reversing rotation loads may lead to excessive friction and wear.
U.S. Pat. No. 1,948,827 utilises a rotation arm mounted pivotally off of the reciprocator with a pivot axis perpendicular to the reciprocator axis passing through point X. The tip of this rotation arm is linked through a short connecting rod to an eccentric shaft rotating at twice the speed of the crankshaft. This produces a rotation restraint system that is closer to ideal than a universal joint. This mechanism also has extra complexity due to the auxiliary shaft that must be indexed to the crankshaft rotation at twice its speed. Such a rotation restraint method may be even more difficult to incorporate within an axial piston machine in which the cylinder cluster is spinning such as described in U.S. Pat. No. 3,654,906 (Airas).
U.S. Pat. No. 4,235,116 shown in FIG. 1 and FIG. 2, describes a rotation restraint method that utilises two restraint gimbals 60, 70. The two gimbals are linked together at a point 64.
U.S. Pat. No. 4,235,116 suggests that the link 64 between the two gimbals 60, 70 should consist of a ball joint or universal joint. However ball joints are generally not suited to high speed reciprocating motion under high-load conditions. Such conditions can lead to rapid wear and short life. This may be further exacerbated by the high swash angle and hence larger range of ball joint motion. Utilising a conventional universal joint having two perpendicular rotational degrees of freedom at point 64 (such as are commonly found in automotive drive trains) may not have sufficient rotational degrees of freedom to prevent the joint from becoming over-constrained and locking up. Over-restraint of this joint may mean that the joint, and by extension the rotation restraint mechanism, cannot move freely as required.
One mechanism layout suggested by U.S. Pat. No. 4,235,116 has the restraint gimbals 60, 70 formed from large rings or portions of rings 63, 73 that fit around the outside of the reciprocator 34 and connecting rods 41. This means that the casings surrounding the reciprocator may need to be increased in diameter in order to accommodate the gimbals, potentially increasing the overall size and weight of the axial piston machine. Because this gimbal structure connects the two gimbal hinge bearings 38 or 71, and the gimbal tip pivot 64 with what are in effect bowed beams, such gimbals may generally need to be made much heavier than if they were comprised of more structurally efficient straight beams in order to have sufficient rigidity to bear the rotation restraint and inertial forces encountered in high speed operation. This greater mass may lead to still higher inertial loadings and a requirement for heavier gimbal hinge and gimbal linking bearings, and can greatly increase the loads on the reciprocator bearings 32. The higher bearing loads created by heavy gimbals leads to increased frictional losses and may limit the maximum speed and power of the machine.
In order to balance the inertial forces created by the motion of the gimbals U.S. Pat. No. 4,235,116 teaches that the gimbals 60, 70 be balanced by mass 76 and 66 such that their respective centres of mass are located on the axes of their respective hinge bearings 38, 71, to allow for the balancing of inertial forces in the machine. This is a restrictive solution to the problem of balancing the gimbals. The practicality of employing these solutions for the balancing of the gimbals in a high speed axial piston engine may be severely hampered by the extra weight and bulk of the gimbals and other elements that can make it difficult to package the mechanism compactly. The extents of the gimbals on the opposite side of the hinge axis from the gimbal linking pivot joints may be difficult to accommodate without interfering with the desired positions of other components. At high speeds the greater rotational inertias and masses of these balanced gimbals may also lead to very high inertial loads that are transmitted through the gimbal and reciprocator bearings. These gimbal inertia-induced bearing loads may be impractically high, limiting the maximum operating speed of the machine, limiting life and leading to increased frictional losses. Undesirable gyroscopic forces may also be established.
It is accordingly an object of this invention to provide a rotation restraint mechanism for z crank axial piston machines that may offer a number of improvements to some or all of the disadvantages that are outlined by reference to the prior art discussed above or to at least offer the public a useful choice.